Control unit of lockup clutch for vehicle

i

DIRECT TAKE-OFF CLUTCH CONTROL DEVICE

VEHICLES AND ITS ORDERING PROCEDURE

BACKGROUND OF THE INVENTION

1. Field of Invention

This invention relates to a device for controlling a lock-up clutch of a vehicle and a method for controlling the latter and, more particularly, to a device for controlling a clutch with direct drive of a vehicle, capable of extending the duration of execution of a fuel cut command, and relates to a control method therefor.

2. Description of the Art Concerned

There is a technique for increasing the fuel cut-off time by a slip commander of a lockup clutch when a vehicle is decelerating, so as to improve fuel consumption, and this technique is described by example, in Japanese Unexamined Patent Application No. 11-257484. During a decelerating slip, in which the lockup clutch is slid-controlled, when the vehicle is decelerating, since it is possible to keep the engine speed at substantially the same speed as the turbine rotation speed, it is possible to maintain the speed of rotation of the engine at a higher value compared with the case of disengaging the direct-drive clutch and it is possible to increase the duration of the fuel cut-off.

However, in the above-mentioned technique, when the vehicle is decelerating and a gear ratio of an automatic transmission is high, because the fuel cut-off control is completed earlier than in the case of which is in a low gear ratio, there is a margin to obtain an improvement in the sense of improving fuel consumption. That is, a condition of execution of the fuel cut command is imposed a lower limit from the engine rotational speed. Further, the rotational speed of the turbine, for a vehicle speed, is lower when in a high speed ratio than it is in a low speed ratio. Even when the direct-drive clutch is slip-controlled, the engine rotational speed is maintained at practically the same rotational speed as the turbine rotational speed, so that the problem arises that the engine cut-off time. Fuel becomes shorter at high gear than it is at low gear.

Further, at the time of deceleration, by initially setting the gear ratio low, the fuel cut-off time is increased; however, the problem arises that an engine brake is applied too strongly and a feeling of discomfort is generated for the driver, so that it is difficult to put this into practice.

The reason why a problem is created is that since the speed •. ^r speed of the turbine and the rotational speed of the engine become large at a downshift moment when the vehicle is at high speed, the friction torque of the engine becomes high. In addition, the moment of inertia is added, which is accompanied by an increase in the rotational speed of the engine and the engine brake torque is amplified by the increase in the transmission ratio.

SUMMARY OF THE INVENTION

To solve the above-mentioned problem, an object of the invention is to provide a device for controlling a direct-drive clutch, which makes it possible to increase the duration of execution of a fuel cut-off command, and a process for its order.

According to a first aspect of the invention, there is proposed a device for controlling a direct-drive clutch of a vehicle. The control device has a fuel cut-off control element which features a hydraulic power transmission with a direct-drive clutch that directly connects an engine and an automatic transmission, stopping the fuel supply to the engine when the engine speed is reached. vehicle is being reduced, and restarting the fuel supply to the engine at a time of judgment that the engine rotational speed becomes a predetermined fuel cut-off end rotational speed, a slip control element which performs a slip control of the lockup clutch when the vehicle is decelerating and a maneuver control element which downshifts the automatic transmission at a time of judgment that the engine rotational speed becomes a downshift rotational speed which is greater, by a determined amount, than the end of fuel cut-off rotational speed, during execution of the slip control by the slip control part.

According to one embodiment of the invention, this device for controlling the direct-drive clutch of a vehicle, comprising a hydraulic power transmission with a direct-drive clutch which directly connects the engine and the automatic transmission, comprises:

fuel cut control means for stopping the supply of fuel to the engine when the vehicle is decelerated, and restarting the supply of fuel to the engine at the judgment time when the rotational speed of the engine becomes a speed of predetermined fuel cut-off end rotation;

slip control means for performing a slip control of the lockup clutch at the time the vehicle is decelerating; and maneuver control means for lowering the gear of the automatic transmission at a judging moment when the rotational speed of the engine and / or the rotational speed of the input shaft of the automatic transmission becomes a speed of. lower gear downward rotation, by a predetermined value, at the end of fuel cutoff rotational speed, during execution of the slip control by the slip control means.

According to the above-mentioned aspects, at the moment at which the engine speed of rotation becomes greater, by a predetermined value, than the speed of rotation at the end of fuel cut-off and this, during the slip control of the clutch at direct drive and when the vehicle is decelerating, the automatic transmission downshift maneuver is performed. Since an input shaft rotational speed of the automatic transmission increases in correspondence with the downshift of the automatic transmission, it is possible to prevent the engine rotational speed from being reduced. As a result, it is possible to increase the time required for the engine rotational speed to reach the fuel cut-off end rotational speed, and it is possible to increase the fuel cut-off time. Further, since the downshift is performed immediately before the engine rotational speed reaches the fuel cut-off end rotational speed, it is possible to limit the detrimental effect, which gives an uncomfortable feeling to the driver. due to the increase in the deceleration of the vehicle caused by an increase in the engine friction torque and the inertia torque.

Further, the slip control part can change the slip degree of the lockup clutch to limit the increase in engine rotation speed in correspondence with the downshift operation of the automatic transmission.

Since the lockup clutch is slid to limit the increase in engine rotational speed caused by downshifting at a time of performing the automatic transmission downshifting, it is possible limit the increase in engine friction torque and inertia torque, and it is possible to properly restrict the increase in vehicle deceleration.

Further, the slip control part can slip the lockup clutch, to prevent the engine rotation speed from increasing in correspondence with the downshift of the automatic transmission.

Since the lockup clutch is slid to prevent the engine rotational speed from being increased, it is possible to properly limit the increase in engine friction torque and inertia torque.

Further, the slip control part can slip the lockup clutch, so as to maintain the engine rotational speed during the period following the downshift of the automatic transmission and until the speed of rotation of the input shaft of the automatic transmission reaches the speed of rotation predetermined for the speed of rotation of the engine, immediately before lowering of the automatic transmission ratio.

Since during the period following downshifting by the automatic transmission and until the speed of rotation of the input shaft of the automatic transmission reaches the predetermined speed of rotation, the speed engine rotation speed is maintained at the engine speed that was immediately before the automatic transmission downshift, it is possible to correctly limit both the increase in the inertia torque and the reduction in the engine speed.

According to a second aspect of the invention, there is provided a device for controlling a direct-drive clutch of a vehicle. The control device has a slip control element, which has a direct-drive clutch directly connecting an engine and an automatic transmission, and which increases the fuel cut-off time required from the start of the fuel cut to which the power is supplied. fuel to the engine is stopped and until an engine rotational speed reaches a fuel cut-off end rotational speed, by performing a slip control on the lockup clutch when the vehicle is decelerating, a transmission shift control element that increases fuel cut-off time by downshifting the automatic transmission, immediately before that the engine rotational speed reaches the fuel cut-off end rotational speed.

According to the second aspect, since the speed of rotation of the input shaft of the automatic transmission becomes important if the automatic transmission ratio is lowered immediately before the speed of rotation of the engine reaches the speed of rotation at the end of the fuel cut, it is possible to limit the reduction in engine speed. As a result, it is possible to increase the time required until the engine rotational speed reaches the fuel cut end rotational speed, and it is possible to increase the fuel cut time. Further, since the downshift is made immediately before the engine rotational speed reaches the fuel cut end rotational speed, it is possible to limit the detrimental effect of an uncomfortable feeling given to the driver. due to the increase in the deceleration of the vehicle in accordance with the increase in the friction torque of the engine and the torque of inertia.

According to a third aspect of the invention, there is provided a method of controlling a lockup clutch of a vehicle having a hydraulic power transmission provided with a lockup clutch, which directly connects an engine and a transmission. automatic by stopping the fuel supply to the engine, when the vehicle speed is reduced, and resuming supplying fuel to the engine at a time of judgment that an engine rotational speed becomes a predetermined fuel cut-off end rotational speed. The method further includes performing a slip control on the lockup clutch when the vehicle is decelerating, downshifting the automatic transmission at a time of judgment when the engine rotational speed. becomes a downshift rotational speed which is greater, by a predetermined amount, than the fuel cut-off end rotational speed, during execution of the slip control by the slip control part.

According to a fourth aspect of the invention, there is provided a method of controlling a lockup clutch of a vehicle having a lockup clutch directly connecting an engine and an automatic transmission. The control method includes increasing the fuel cutoff time required from the start of the fuel cutoff at which fuel supply to the engine is stopped, and until a rotational speed of the engine reaches a speed. rotation end of fuel cut, by performing a slip control of the direct-drive clutch, when the vehicle is decelerated, and increasing the automatic transmission downshift fuel cut time immediately before the engine rotational speed reaches the fuel cut end rotational speed.

According to the third and fourth aspects, since the speed of rotation of the input shaft of the automatic transmission becomes high, due to the lowering of the ratio of the automatic transmission, immediately before the speed of rotation of the engine n 'reaches the end of fuel cut-off speed, it is possible to limit the reduction in engine speed. As a result, it is possible to increase the time required until the engine rotational speed reaches the fuel cut end rotational speed, and it is possible to increase the fuel cut time.

BRIEF DESCRIPTION OF THE DRAWINGS

Fig. 1 is a view showing a power transmission device for a vehicle to which a control device of a lockup clutch according to an embodiment of the invention is applied;

Fig. 2 is a table describing the relationship between a combination of operations between a first solenoid valve and a second solenoid valve and a gear ratio change obtained in this way, in an automatic transmission provided with a torque converter having a clutch clutch. direct;

Fig. 3 is a view illustrating a main structure of a hydraulic control circuit;

Fig. 4 is a view illustrating an output characteristic of a linear solenoid valve shown in FIG. 3;

Fig. 5 is a view showing a property of a slip control valve provided in the hydraulic control circuit of FIG. 3;

Fig. 6 is a view showing a change as a function of time between an engine rotational speed and a turbine rotational speed, obtained by a control operation of an electronic controller and a clutch or clutch pressure. direct drive; and

Fig. 7 is a flowchart showing the content of the control of the lockup clutch.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

A description will be given in detail below concerning an embodiment according to the invention, with reference to the accompanying drawings.

Fig. 1 is a view illustrating a main part of a power transmission device for a vehicle to which an embodiment according to the invention is applied. With reference to FIG. 1, the power of a motor 10 is transmitted to a differential gear and a drive wheel (which are not shown) via an automatic transmission 14 of the "gear ratio" type, formed by a torque converter 12 with a direct-drive clutch, three sets of sets of planetary gear wheels and the like.

The torque converter 12 mentioned above is provided with a pump wheel 18, fitted with blades, connected to a crankshaft 16 of the engine 10, a turbine wheel 22, provided with blades, fixed on the shaft of 'input 20 to the automatic transmission 14 and driven in rotation by the oil coming from the pump wheel 18, provided with blades, a massive stator wheel 28 fixed to a housing 26 corresponding to a non-rotating member, via a one-way clutch 24 and a direct-drive clutch 32 connected to the input shaft 20 via a damper 30. When a hydraulic pressure prevailing in a hydraulic chamber 33 on the clutch side is increased to a value greater than that in the hydraulic chamber 35 on the clutch side inside the torque converter 12, the clutch engagement direct 32 is disengaged, so that the torque is transmitted with an amplification factor which corresponds to the ratio between the input and output rotational speed of the torque converter 12. However, when the hydraulic pressure prevailing inside the hydraulic chamber 35 on the clutch side is made greater than the hydraulic chamber 33 on the clutch side, the lockup clutch 32 is engaged, so that the input member and the output member of the torque converter 12, that is to say the crankshaft 14 and the input shaft 20, have entered a direct drive state.

The automatic transmission 14 is provided with three sets of single pinion type planetary gear devices 34, 36 and 38 which are mounted coaxially, the input shaft 20 mentioned above and a countershaft (an output shaft 40 which transmits power between an output gear 39 rotating together with a ring gear of the planetary gear apparatus 38 and the differential gear apparatus. Some of the constituent elements of the planetary gear units 34, 36 and 38 are not only integrally connected to each other but are also selectively connected to each other by three clutches C0, C1 and C2. Further, part of the component parts of the planetary gear apparatuses 34, 36 and 38 are selectively connected to the housing 26 by four brakes B0, B1, B2 and B3, and part of the component parts are constructed so as to be engaged. with each other or with the housing 26 according to the direction of rotation of the latter by three unidirectional clutches F0, Fl and F2.

The clutches C0, Cl and C2 and the brakes B0, Bl B2 and B3 are constituted, for example, by a multi-disc type clutch, and a band brake provided with a band or with two bands having rolling directions opposites, or the like, and are constructed to be operated, respectively, by hydraulic actuators. Then, the operations of the hydraulic actuators are respectively controlled by an electronic controller 42 to be mentioned below, so that a gear change speed, with four forward gears and one reverse gear, respectively, having ratios of different transmission I (= speed of rotation of input shaft 20 / speed of rotation of counter shaft 40), can be obtained as shown in Fig. 2. In Fig. 2, the indications of 1^er, 2^eme, 3^eme_et O / D (Overdrive) respectively denote a first gear ratio, a second gear ratio, a third gear ratio and a fourth gear ratio, in forward direction and the above mentioned gear change is reduced sequentially, from the first to the fourth gear. In this case, since the torque converter 12 and the automatic transmission 14 are constructed symmetrically with respect to an axis, Fig. 1 shows them in a way which omits to illustrate the underside of an axis of rotation of the input shaft 20 and an upper side of an axis of rotation of the counter shaft 40.

Further, a hydraulic control circuit 44 is provided with a hydraulic control circuit controlling the gear shift transmission to control the speed ratios of the automatic transmission 14 and a hydraulic control circuit controlling the lockup clutch, in order to to control the engagement operation of the direct-drive clutch 32. The hydraulic control circuit controlling the gear change transmission is provided with a first solenoid valve 46 and a second solenoid valve 48, which are respectively activated and deactivated by a solenoid No. 1 and a solenoid No. 2, as it is well known. In addition, the construction is such that the clutch and the brake are selectively operated as shown in Fig. 2 from the combination of operation between the first solenoid valve 46 and a second solenoid valve 48, so that any gear ratio among the first or fourth gear ratio, mentioned above, is established.

Further, the hydraulic control circuit controlling the lockup clutch is provided, for example, as shown in Fig. 3, with a third electromagnetic valve 50 which is activated and shut off by an electromagnetic switching solenoid 49, so to generate a switching signal pressure Psw, a clutch switch valve 52 which is switched to a side-disengaged position so that the clutch for direct-drive 32 goes to the disengaged state and to a side-engaged position for shifting the direct-drive clutch 32 in the engaged state corresponding to the switching signal pressure Psw, a linear solenoid valve 54 which generates a slip control signal pressure PSLU corresponding to a drive electric current USLU supplied from the electronic controller 42, and a slip control valve 56 which adjusts the signal pressure difference of switching ΔΡ between the engagement side hydraulic chamber 35 and the disengaged side hydraulic chamber 33 serving as engagement pressure of the lockup clutch 32, corresponding to the PLSU slip control signal pressure produced from the linear solenoid valve 54, and controls the amount of slip of the lockup clutch 32.

In Figure 3 mentioned above, a pump 60, for sucking and supplying a working fluid pressure returning to this reservoir (not shown) via a trap 58, is constructed so as to be rotated by the motor 10. The pressure of the working fluid supplied by the pump 60 is adapted to be adjusted to a first piping pressure P1 by a first pressure adjustment valve 62 of the overflow type. This first pressure adjustment valve 62 generates the first piping pressure Pli which is correspondingly increased to a throttle pressure produced by a throttle valve opening detection valve (not shown), and produced via a first passage of piping oil 64. A second pressure adjustment valve 66 is an overflow type pressure adjustment valve and generates a second line pressure P12 corresponding to an output torque of the engine 10 by adjusting a pressure of the working fluid output from it. the first pressure adjustment valve 62 according to the throttling pressure mentioned above. A third pressure adjustment valve 68 is a pressure reducing valve in which the first piping pressure P1 is an original pressure, and produces a third piping pressure P13 of fixed value. Further, a manual valve 70 generates a pressure PR in the range R, when a maneuvering actuating lever 196 is in the range R. In addition, an OR valve 72 selects and produces a pressure PB2 for the operation of the brake B2 put in direct drive in the second higher gear ratio or higher or the in the pressure range R, mentioned above, which is more high.

The above-mentioned clutch changeover valve 52 is provided with a clutch side port 80 which is communicated with the clutch side hydraulic chamber 33, a clutch side port 82 which is communicated with the chamber. clutch side hydraulic 35, an inlet port 84 to which the second line pressure P12 is supplied, a first discharge port 86 from which the working fluid in the hydraulic chamber on the clutch side 35 is discharged when the clutch for engagement 32 is disengaged, a second discharge port 88 from which the fluid work in the clutch side hydraulic chamber 33 is unloaded when the clutch for engagement 32 is engaged, a supply port 90 to which a portion of the working fluid discharged from the second pressure adjustment valve 66 is supplied for cooling during a period of engagement of the clutch for engagement 32, a valve member a coil 92 which switches a connection state of the ports, a spring 94 which moves the coil valve member 92 to a stop side position, a piston 96 arranged to be capable of abutting an end portion in one side of the spring 94 of the spool valve member 92, a hydraulic chamber 98 provided between end surfaces of the valve member coil 92 and piston 96 for applying pressure PR in the range R, to the end surfaces, a hydraulic chamber 100 which receives the first pipe pressure Pli applied to the end surface of piston 96, and a hydraulic chamber 102 which applies a switching signal pressure Psw produced by the third solenoid valve 50 to the end surface of the spool valve member 92 and receives the switching signal pressure Psw to generate a thrust in the direction. from a walking side position.

The third solenoid valve 50 is constructed so that a spherical valve member or shutter cuts off communication between the hydraulic chamber 102 and the OR valve 72 in an un-excited state (off state) and the hydraulic chamber 102 is set at a pressure of drain, however in an energized state (run state), communication of the hydraulic chamber 102 with the OR valve 72 is established to apply the switching signal pressure Psw to the hydraulic chamber 102. Correspondingly, when the third solenoid valve 50 is in the off state, the switching signal pressure Psw produced by the third solenoid valve 50 is not applied to the hydraulic chamber 102 and the spool valve member 92 is turned off. placed in the off-side position according to an actuating force exerted by the spring 94 and the first pli pipe pressure applied to the hydraulic chamber 100. Therefore, the inlet port 84 and the clutch side port 80 and the clutch side port 82 and the first discharge port 86 are respectively communicated with each other. Correspondingly, a hydraulic pressure Poff, in the clutch side hydraulic chamber 33 is made higher than a hydraulic pressure Pon inside the clutch side hydraulic chamber 35, so that the clutch for switching on. socket 32 ​​is disengaged. Simultaneously, the working oil in the clutch side hydraulic chamber 35 is discharged to the drain via the above-mentioned first discharge port 86, an oil cooler 104 and a check valve 106.

On the contrary, when the third solenoid valve 500 is in the on state, the switching signal pressure Psw produced by the third solenoid valve 50 is applied to the hydraulic chamber 102 and the spool valve member 92 is set to the side position. works against the actuating force exerted by the spring 94 and the first hydraulic pressure of pli piping applied to the hydraulic chamber 100, so that the inlet port 84 and the clutch side port 82, the clutch side port 80 and the second discharge port 88 and the supply port 90 and the first discharge port 86 are set. in respective communication. Likewise, the hydraulic pressure Pon in the clutch side hydraulic chamber 35 is made higher than the hydraulic pressure Poff in the clutch side hydraulic chamber 33 and the clutch for engagement 32 is engaged and simultaneously the Working fluid in the clutch side hydraulic chamber 33 is discharged to the drain via the second discharge port 88 and the slip control valve 56.

The above-mentioned linear solenoid valve 54 is a pressure reducing valve in which the set third line pressure P13 generated in the third pressure adjusting valve 68 is an original pressure, generates a signal pressure of slip control PSLU which increases correspondingly to a control electric current PSLU produced by the electronic control device 42 as shown in Fig. 4, and applies the slip control signal pressure PSLU to the slip control valve 56. The linear solenoid valve 54 is provided with a supply port 110 to which it supplies the third pipe pressure P13, a port of. exit 112 which provides the. PSLU slip control signal pressure of a coil valve member 114 which opens and closes them, of a spring 115 which actuates the coil valve member 114 in the opening direction of the valve, a spring 116 which actuates the coil valve member 114 in the direction of the valve opening by a lower thrust than that of the spring 115, a slip control electromagnetic solenoid 118 which operates the coil valve member 114 in the direction of valve opening according to the ISLU control electric current, and a hydraulic chamber 120 which receives a reaction pressure (the pressure slip control signal PSLU) to generate the thrust in the opening direction of the valve, in the spool valve member 114, and the coil valve member 114 is operated such that the actuating force produced in the opening direction of the valve obtained by the electromagnetic solenoid 118 and the spring 116 balances with the actuating force produced. in the closing direction of the valve, obtained by the spring 115 and the reaction pressure.

The slip control valve 56 is provided with a line pressure port 130 to which the second line pressure PI2 is supplied, a receiving port 132 which receives the working fluid in the unloaded clutch side hydraulic chamber 33. from the second discharge port 88, a drain port 134 for discharging the working fluid received in the receiving port 132, a spool valve member 136 which is provided so as to be movable in the direction to a first position (a right-hand position in Fig. 3) by increasing a pipe pressure difference AP (= Pon - Poff) between the clutch side hydraulic chamber 35 and the clutch side hydraulic chamber 33) by a communication between the receiving port 132 and the drain port 134 to discharge the working fluid present in the chamber hydraulic clutch side 33 and to a second position (a left position in Figure 3) reducing the value Δ P mentioned above by establishing communication between the reception port 132 and the pipe pressure port 130 so as to provide the second piping pressure P12 in the hydraulic chamber on the clutch side 33, a piston 138 arranged to be able to abut against the coil valve member 1S6 to actuate the coil valve member 136 to the first position, a signal pressure hydraulic chamber 140 which receives the slip control signal pressure PSLU to apply the slip control signal pressure PSLU to the piston 138 and the spool valve member 136 to respectively generate the thrusts in directions spacing from each other in piston 138 and spool valve member 136, a hydraulic chamber 142 which receives the hydraulic pressure Poff to apply the hydraulic pressure Poff inside the clutch side hydraulic chamber 33 to the piston 13S to generate the thrust in a direction of travel of the spool valve member 136 in direction of the first position in the piston 138, a hydraulic chamber 144 which receives the hydraulic pressure Pon to apply the hydraulic piping pressure Pon inside the clutch side hydraulic chamber 35 to the spool valve member 136 to generate the thrust in a direction towards the second position in the spool valve member 136 and a spring 146 is housed in the hydraulic signal pressure chamber 140 to actuate the spool valve member 136 in the forward direction. second position.

In this case, a first zone 148 and a second zone 150 having cross-sectional areas A1 and A2 which sequentially become smaller from the side of the hydraulic chamber 142, are formed in the piston 138, and a third zone 152 having a cross-sectional area A3, a fourth area 154 having a cross-sectional area A4 whose cross-sectional area is smaller than that of A3 and the same as that of area A1, and a fifth area 156 having a cross-sectional area A5 which is the same as the area A1, are formed in the spool valve member 136, from the side of the hydraulic signal pressure chamber 140. The areas of Cross-sectional areas are in a relation A3> Al (= A4 = A5)> A2. Correspondingly, when the clutch switch valve 52 is on, the slip control signal pressure PSLU is comparatively low, and a relationship shown in a formula (1) is established and the piston 138 abuts the spool valve member 136 to be fully applied to each other, so that a pressure difference Δ P of a value corresponding to the slip control signal pressure PSLU is formed. At this time, the pressure difference AP changes relatively gradually at an inclination [(A3 - A2) / A1] based on a formula (2) with respect to the slip control signal pressure PSLU. In this case, in formula (2), Fs is an actuating force of the spring 146.

Al 'Poff> A2' PSLU (1)

ΔΡ - Pon - Poff = [(A3 - A2) / A1] PSLU - Fs / Al (2)

However, when the slip control signal pressure PSLU exceeds a value greater than a predetermined PA value, a relationship represented by a formula (3) is established. The predetermined value PA is a value which is previously determined so that a range of variation zïPslip of the pressure difference ΔΡ of sufficient value necessary for the slip control of the lockup clutch 32 can be obtained, and the areas of respective cross-section and the like are set so that the relationship shown in formula (3) is established when the slip control signal pressure PSLU assumes the value PA. Correspondingly, the piston 138 and the spool valve member 136 move away from each other, and the spool valve member 136 is operated so that the formula (4) is set. However, since the slip control valve 56 is constructed so that the receiving port 132 thereof and the drain port 134 are in communication with each other in a state that the When the spool valve element 136 is operated so that formula (4) is set, the hydraulic pressure Poff inside the clutch side hydraulic chamber 33 is further reduced to become atmospheric pressure, a relation Δ P = Pon is established and a complete clutch is established. A solid line in Fig. 5 shows a change property of the pressure difference Δ P obtained by actuation of the slip control valve 56 constructed in the manner described above with respect to the control signal pressure of PS LU slip.

Al 'Poff <A2' PSLU (3)

A3 = PSLU = A4Pon + Fs (4)

Further, as shown in Fig. 5, when the slip control signal pressure PSLU becomes low to take the value PB in which a formula (5) is established, the relation of the pressure difference ΔΡ = 0 is established , so that the lockup clutch 32 is disengaged despite the switch valve 52 being in the on state.

A3 'Pon> A3' PSLU (5)

Referring again to Fig. 1, the electronic controller 178 for execution by the engine of a fuel injection command, a fuel cut command, a timing command. of the picture and the like, in which an amount of injected fuel is controlled by a fuel injector (not shown) corresponding to an amount of intake air, is installed in a vehicle. In this fuel cut-off command, when the engine speed NE of rotation becomes equal to or greater than a pre-set fuel cut-off rotational speed Ncut at the time of coasting, deceleration, in which the opening of the valve d 'TA throttling is equal to or less than a predetermined value, close to 0, the fuel supply to the engine 10 is stopped, for example by closing the fuel injector. This cessation of fuel supply is aimed at optimizing fuel consumption.

The electronic controller 42 is a so-called microcomputer formed by CPU 182, ROM 184, RAM 186, interface (not shown) and the like. It is designed so that a signal which expresses a degree of opening of the throttle valve TA, a signal expressing the rotational speed NE of the motor (i.e., a rotational speed on the input side of the l direct drive clutch 32), a signal expressing an input shaft rotational speed Nin (a turbine rotational speed NT, i.e. a rotational speed of the output side of the clutch at direct drive 32), a signal expressing a speed of rotation of the output shaft Nout corresponding to a vehicle speed V, and a signal expressing an operating position Ps of the operating lever 196 are respectively supplied to the electronic control device 42 from a sensor of throttle 188 provided in an intake pipe of the engine 10 to detect an opening of a throttle valve 187 opened and closed by actuation of an accelerator pedal (not shown), an engine rotational speed sensor 190 which detects the rotational speed of the engine 10, an input shaft rotational sensor 192 which detects a rotational speed of the input shaft 20 in the automatic transmission 14, a countershaft rotation sensor 194 which detects the rotational speed of the countershaft 40 in the automatic transmission 14 and an operating position sensor 198 for detecting the operating position of the operating lever 196, that is, one of the operating ranges L, S, D, N, R and P. The CPU 182 of the above-mentioned electronic controller 42 processes the input signals according to a program previously stored in the ROM 184 while using a temporary storage function of the RAM 186, executes a maneuver transmission command of the automatic transmission 14 and a clutch control for the lockup clutch 32 according to a main program (not shown), and controls the first solenoid valve 46, the second solenoid valve 48, the third solenoid valve 50 and the linear solenoid valve 54, respectively.

In the above-mentioned shunting transmission control, a shunting transmission graph corresponding to an actual shunting transmission ratio is selected based on a plurality of types of shunting transmission graphs previously stored in ROM 184. and the maneuvering transmission ratio is determined on the basis of the traveling state of the vehicle, for example, the opening of the throttle valve TA and the speed of the vehicle, calculated from an output shaft rotation speed Nout, according to the maneuver transmission graph, and the first solenoid valve 46 and the second solenoid valve 48 are driven so that the maneuver transmission ratio ratio is obtained, so that the operations of the clutches C0, Cl and C2 and of the brakes B0, Bl, B2 and B3 in the automatic transmission 14 are controlled, and any one of the engagement states of the four forward gears be established.

The clutch control of the lockup clutch 32 mentioned above is constructed to be performed, for example, during movement in the second gear ratio, the third gear ratio and the fourth gear ratio. In clutch control according to the relationship previously stored in ROM 184 as requested by a clutch control program (not shown), it is determined whether the lockup clutch 32 is in the disengage zone, the Slip control zone or the clutch zone according to the appearance of the vehicle, so that the speed of rotation of the output shaft (vehicle speed) Nout and the opening of the throttle valve TA. The third solenoid valve 50 and the linear solenoid valve 54 are driven based on the result of judgment so that the operation of the lockup clutch 32 is controlled.

Further, the electronic controller 42 performs a so-called deceleration slip control which slidably controls the lockup clutch 32 during execution of a fuel cut command at a displacement moment. coasting, decelerating, mentioned above and executes a downshift transmission command of the automatic transmission 14 during the decelerating slip, thus lengthening the fuel cut-off time. Fig. 6 is a view for explaining the control operation according to the above-mentioned control process and illustrates the change over time of the rotational speed NE of the engine and the rotational speed NT of the turbine when the device electric control 42 executes the above mentioned control and when the engagement pressure of the lockup clutch 32.

With reference to FIG. 6, the electronic controller 42 controls the slip of the lockup clutch 32 by adjusting the clutch pressure of the lockup clutch 32 to a predetermined value pressure. A phase of the deceleration slip control is separated into five steps which are called, respectively, steps, and the respective steps are shown in a top step in FIG. 6.

A step 1 corresponds to a state of adjusting the clutch pressure of the lockup clutch 32, so that the degree of slip of the lockup clutch 32 becomes a predetermined setpoint slip value when the deceleration slip command is in execution condition. Correspondingly to the deceleration, the rotational speed NT of the turbine, and the rotational speed NA of the engine decrease to a level, for example, from 1800 revolutions / minute to 900 revolutions / minute, as shown in FIG. 6. In the fuel cutoff control, a fuel cut end rotational speed Nret is previously determined, and when the engine rotational speed NE becomes equal to or less than the fuel cut end rotational speed, Nret ( about 700 rpm in Fig. 6), the fuel cut control is ceased and the fuel supply to engine 10 is restarted.

Correspondingly, the downshift of the automatic transmission 14 is performed at a stage at which the turbine rotational speed NT becomes a predetermined downshift judgment rotational speed NT1 which is greater than the rotational speed. Nret of end of fuel cut, that is to say at a stage at which the engine speed of rotation NE becomes greater than the speed of rotation Nret of end of fuel cut (a step 2), this by a predetermined amount. Then, the turbine rotational speed NT is increased, the reduction in the engine rotational speed NE is limited, and the fuel cut-off time is increased. That is, a reduction in the engine speed NE of rotation is limited, so that it is possible to increase the time required until the engine speed NE reaches the speed of Nret fuel cut-off end rotation.

In this case, the NT1 downgrade judgment rotational speed is generally set to a value at which the predetermined amount becomes between a few tens of revolutions per minute and one hundred revolutions per minute and a few tens of revolutions per minute, although the NT1 speed is different in the respective speed ratios. This value is set prior to each gear and is stored in ROM 184 of electronic controller 42.

In this case, the judgment of the execution of the downshift can use the engine speed NE instead of using the turbine speed NT or the downshift can be performed when the any one of the turbine rotational speeds NT or the rotational speed NE reaches the downshift judgment rotational speed.

Further, at this time, the engine rotational speed NE is increasing corresponding to the turbine rotational speed NT increase. The above-mentioned increase in engine rotational speed NE causes increased friction in the engine and inertia torque and increased vehicle speed reduction. In step 3 and step 2, to limit the increase in vehicle speed reduction mentioned above, the lockup clutch 32 is slidably controlled so as to restrict the increase in rotational speed NE of the engine corresponding to the downshift of the automatic transmission 14. In particular, a clutch pressure of the clutch for the engagement 32 is adjusted, as shown at a lower step in FIG. 6, so that the lockup clutch 32 is slid to maintain the engine speed NE in the engine speed immediately before downshifting.

The turbine rotational speed NT is reduced in this state, and when a difference between the turbine rotational speed NT and the engine rotational speed NE occurs, i.e. an amount of slip of the lockup clutch 32 becomes small to achieve a set amount of slip, the lockup clutch pressure 32 is adjusted, so that the slip amount of the lockup clutch 32 becomes a predetermined setpoint slip amount, in the same manner as that of gear 1 (bare step 4).

Further, when the engine rotational speed NE, or the turbine rotational speed NT becomes equal to or less than a final rotational speed with decelerating slip, the slip control of the lockup clutch 32 is completed, and this direct-drive clutch 32 is disengaged (step 5).

Fig. 7 is a view showing an example of a particular processing flow for performing the control operation shown in FIG. 6. With reference to FIG. 7, firstly, at a step S11, it is judged whether or not the condition of execution of the deceleration slip control is established. If the condition for executing the deceleration slip command is established (YES to S11), that is to say if step 1 mentioned above is established, a command of the setpoint slip degree to command the engagement pressure (the clutch pressure) of the lockup clutch 32 is then executed in a step S12 so that the slip degree of the lockup clutch 32 becomes a setpoint slip amount Predetermined TNSLP.

Next, in a step S13, it is judged whether or not the speed ratio is a minimum speed ratio performing slip in deceleration. For the minimum speed ratio performing slip in deceleration for example, a lowest speed ratio, capable of shifting the motor 10 to a driven stall at the time of coasting and deceleration, is fixed. When the ratio is judged not to be the minimum deceleration slip execution ratio (NO to S13), that is, it is judged that the deceleration slip control can be executed even when a downshift is carried out on the automatic transmission 14, then a step S14 is judged whether or not the speed of rotation of the turbine is greater than the speed of rotation for judging the downshift, NT1.

When the turbine NT rotational speed becomes equal to or lower, the NT1 downshift judgment rotational speed (answer NO to S14), the gear changes to the above-mentioned 2 and 3 ratio and then , downshifting from the automatic transmission 14 is performed at a step S15. Further, in step S16, an engine rotational speed TNE is set to limit the increase in the engine rotational speed NE corresponding to downshift, and the clutch pressure is controlled so that the motor rotation speed NE becomes the setpoint rotation speed TNE. The engine setpoint TNE speed is fixed at the engine speed immediately before downshifting, however, it

C: \ WINDOWS \ Temporary Internet Files \ Content.lE5 \ O7NOPN0G \ tfn020085 (French_ $ B! K_ (B (l) .doc - August 29, 2002 - 1 S / 2 <· can be fixed instead of that at the ratio NT1 descent appreciation rotational speed or else may be set at a predetermined value, slightly greater than the end of fuel cutoff rotational speed Nret.

Then, processing returns to step S13 and judging whether or not the transmission ratio after downshifting and the minimum slip execution ratio with deceleration. The processing is returned to step S13 due to restricting the limitation of the engine rotation speed NE as much as possible, repeating the downshift step until the gear becomes the minimum capable gear. execute the slip control in deceleration and the fuel cut-off time has ceased. As a result, the downshift of the automatic transmission may delay the arrival of the engine rotational speed NE to the fuel cut-off end rotational speed Nret. Instead of processing the return to step S13, it is possible to go to a step S17 from step S16, without having to exit downshift.

Further, if it is judged that the gear after the downshift is not the minimum slip execution ratio with deceleration (NOM response to S13), the downshift is executed again in step S13. after the turbine rotational speed NT becomes equal to or less than the downshift judging rotational speed NT1 (result NO to S14), and in step S16, the clutch pressure is controlled so that the engine speed NE becomes the engine setpoint speed TNE.

In this case, by using the rotational speed NT of the turbine, to appreciate the performance of downshifting in step S14, the downshifting of the automatic transmission in step S15 can be performed, even when the engine rotational speed NE is maintained after the engine rotational speed NE becomes the rotational speed greater than the fuel end cutoff rotational speed Nret, this by a predetermined amount.

Furthermore, if it is appreciated that the speed ratio after the downshift is the minimum execution ratio of the slip in deceleration (answer YES to S13), it is judged in a step S17 whether the speed of rotation of the turbine is or is not greater than a value obtained by adding the motor setpoint speed to the TNSLP setpoint slip quantity. In this case, the rotation speed NE of the motor can be used instead of the reference speed TNE of this motor. In addition, if the turbine rotation speed NT becomes equal to or less than an amount obtained by adding the motor setpoint rotation speed TNE to the TNSLP setpoint slip quantity (answer NO to S17), the step is continued. moves at the above-mentioned step 4 and the setpoint slip amount control is executed at a step S18.

Then, in a step S19, it is judged whether or not the rotation speed NE of the motor is greater than the rotation speed NE at the end of slip with deceleration. In the case in which the engine rotation speed is greater than the end of deceleration slip rotation speed NE, it is judged in a step S20 whether the turbine rotation speed NT is greater than the rotation speed NT end of slip with deceleration. Further, if any one of the engine rotation speed NE or the turbine rotation speed NT becomes equal to or greater than the control end rotation speed with deceleration, the process goes to step 5 mentioned above. above and the slip control of the lockup clutch 32 is completed at a step S21.

The embodiment described this time should be considered only to serve as an example of the invention for each aspect and not to limit the invention. The scope of the invention is described, not by the above-mentioned description but by the scope of the claims of a patent. In addition, it is evident that the invention includes all modifications in meanings and fields equivalent to the scope of the claims for a patent.